High efficiency internal combustion engine

ABSTRACT

A method for use in generating power in an internal combustion engine includes controlling a flow of a gas having an initial temperature and an initial pressure to a combustion chamber of the engine by a controller through an intake mechanism to provide a first mass of the gas to the chamber. The combustion chamber is expanded from a minimum volume to a maximum volume in an intake stroke. The maximum volume of the chamber exceeds a maximum compression volume of the chamber. The first mass of the gas in the chamber has a first pressure and a first temperature at the maximum compression volume. The chamber is reduced to the compression volume from the maximum volume in a compression stroke. The compression volume is about a volume of the chamber such that the first mass of the gas at the first pressure and the first temperature in the chamber is less than a maximum mass detonating with a fuel by the end of the compression stroke in a gasoline engine. In the case of a diesel engine, the compression volume is about a volume of the chamber such that the first mass of the gas at the first pressure and the first temperature in the chamber is greater than a minimum mass self-igniting with a fuel by the end of the compression stroke. The expansion of the chamber from the compression volume to the maximum volume during the intake stroke and the compression of the chamber to the compression volume during the compression stroke includes a process which is about adiabatic and about reversible. The gas has a second pressure and a second temperature during the process which are lower than the initial temperature and the initial pressure. A mixture of the gas and the fuel in the chamber is ignited to expand the chamber from the minimum volume to the maximum volume in a power stroke. The chamber is reduced to the minimum volume from the maximum volume in an exhaust stroke and exhaust gas and burned fuel are exhausted resulting from the igniting into a surrounding ambient environment. The chamber is expanded from the minimum volume to the maximum volume in a next intake stroke and the flow of the gas is controlled to provide the first mass of the gas to the chamber during the next intake stroke.

CROSS-REFERENCE TO RELATED APPLICATIONS

This application claims priority from U.S. Provisional Application No.60/940,646, filed May 29, 2007, the entirety of which is herebyincorporated by reference.

TECHNICAL FIELD

This invention relates, in general, to Internal Combustion Engines, andin particular, to systems and methods for generating and using power viaInternal Combustion Engines.

BACKGROUND ART

An Internal Combustion Engine (ICE) is a cyclic device in which thethermal energy of burned fuel is partially converted into mechanicalenergy. The fundamental upper limit for efficiency of such a conversionis set by the efficiency of the reversible Carnot cycle operatingbetween the same limiting hot (T_(H)) and cold (Tc) temperatures as thereal engine, η=1−T_(H)/T_(C) (as is well-known, this statementrepresents one of the equivalent formulations of the Second Law ofThermodynamics). For example, an engine operating between the combustiontemperature of gasoline (T_(H)=2,300 K) and room temperature (T_(C)=300K), the Carnot efficiency η is 87%. Conventional ICEs such as Otto cycleengines have efficiencies at best approaching 40-50% using low octanerated fuels available for consumers. Various attempts have been made toimprove the efficiency of internal combustion engines but much room forincreased efficiency remains.

As is well-known, gasoline-fueled ICE operation can be described as anidealized, cyclic process called the Otto cycle. The Otto cycle isdepicted in FIG. 1A in terms of its different stages (strokes), in FIG.1B on a thermodynamic pressure-volume diagram, and in FIG. 2A as apressure vs. cycle phase diagram.

The Otto cycle stages include:

-   -   1) Intake stroke: taking in an amount of gas (air mixed with        fuel); during this stroke the volume of a cylinder increases by        a factor which is called a compression ratio r; for present-day        automobile ICE, r is equal to about 9-10.5.    -   2) Compression stroke: quickly compressing the mixture to a        considerably higher pressure while raising its temperature. This        compression takes place almost without heat exchange with the        cylinder walls, i.e. near-adiabatically, but at the same time in        an almost-reversible way; in an idealized Otto cycle this is        represented by an adiabatic lower curve on a PV diagram in FIG.        1B. The maximum pressure at point b in FIG. 1B to which the        mixture can be compressed is limited by the condition that it        does not self-ignite (the undesirable phenomenon of        self-ignition is also called detonation or dieseling),    -   3) Ignition: intentionally igniting (e.g., via a spark plug) the        compressed mixture at point b in FIG. 1A, creating a nearly        instantaneous increase in pressure at a fixed volume (i.e., an        isochoric process b-c in FIG. 1B, followed by    -   4) Power stroke: a near-adiabatic and almost-reversible        expansion of combustion products by a volume factor r,        accompanied by the partial transfer of thermal energy of the        burned mixture into mechanical energy; the transfer being        achieved via setting in motion, by the gas pressure, an        appropriate mechanical element such as a moving piston, as shown        in FIG. 1A (e.g., or a rotor in a rotary ICE).    -   5) Exhaust stroke: The moving element (e.g. a piston) pushes out        the used burned mixture through the open exhaust valve while        lowering the pressure back to atmospheric level (another        isochoric process d-a in FIG. 1B)

Thus, there is a need for systems and methods which more efficientlyconvert the internal thermal energy of expanding burned fuel mixture inan internal combustion engine into mechanical energy.

SUMMARY OF THE INVENTION

The present invention provides, in a first aspect, a method for use ingenerating power in an internal combustion engine which includescontrolling a flow of a gas having an initial temperature and an initialpressure to a combustion chamber of the engine by a controller throughan intake mechanism to provide a first mass of the gas to the chamber.The combustion chamber is expanded from a minimum volume to a maximumvolume in an intake stroke. The maximum volume of the chamber exceeds amaximum compression volume of the chamber. The first mass of the gas inthe chamber has a first pressure and a first temperature at the maximumcompression volume. The chamber is reduced to the compression volumefrom the maximum volume in a compression stroke. The compression volumeis about a volume of the chamber such that the first mass of the gas atthe first pressure and the first temperature in the chamber is less thana maximum mass detonating with a fuel by the end of the compressionstroke. The expansion of the chamber from the compression volume to themaximum volume during the intake stroke and the compression of thechamber to the compression volume during the compression stroke includesa process which is about adiabatic and about reversible. The gas has asecond pressure and a second temperature during the process which arelower than the initial temperature and the initial pressure. A mixtureof the gas and the fuel in the chamber is ignited to expand the chamberfrom the minimum volume to the maximum volume in a power stroke. Thechamber is reduced to the minimum volume from the maximum volume in anexhaust stroke and exhaust gas and burned fuel are exhausted resultingfrom the igniting into a surrounding ambient environment. The chamber isexpanded from the minimum volume to the maximum volume in a next intakestroke and the flow of the gas is controlled to provide the first massof the gas to the chamber during the next intake stroke.

The present invention provides, in a second aspect, a method for use ingenerating power in an internal combustion engine which includescontrolling a flow of a gas having an initial temperature and an initialpressure to a combustion chamber of the engine by a controller throughan intake mechanism to provide a first mass of the gas to the chamber.The combustion chamber is expanded from a minimum volume to a maximumvolume in an intake stroke. The maximum volume of the chamber exceeds amaximum compression volume of the chamber. The first mass of the gas inthe chamber has a first pressure and a first temperature at the maximumcompression volume. The chamber is reduced to the compression volumefrom the maximum volume in a compression stroke. The compression volumeis about a volume of the chamber such that the first mass of the gas atthe first pressure and the first temperature in the chamber is more thana minimum mass self-igniting with a fuel by the end of the compressionstroke. The expansion of the chamber from the compression volume to themaximum volume during the intake stroke and the compression of thechamber to the compression volume during the compression stroke includesa process which is about adiabatic and about reversible. The gas has asecond pressure and a second temperature during the process which arelower than the initial temperature and the initial pressure. The mixtureof the gas and the fuel in the chamber is self-ignited to expand thechamber from the minimum volume to the maximum volume in a power stroke.The chamber is reduced to the minimum volume from the maximum volume inan exhaust stroke and exhaust gas and burned fuel are exhaustedresulting from the igniting into a surrounding ambient environment. Thechamber is expanded from the minimum volume to the maximum volume in anext intake stroke and the flow of the gas is controlled to provide thefirst mass of the gas to the chamber during the next intake stroke.

BRIEF DESCRIPTION OF THE DRAWINGS

The subject matter, which is regarded as the invention, is particularlypointed out and distinctly claimed in the claims at the conclusion ofthe specification. The foregoing and other objects, features, andadvantages of the invention will be apparent from the following detaileddescription of preferred embodiments taken in conjunction with theaccompanying drawings in which:

FIG. 1A depicts a cycle of an internal combustion engine;

FIG. 1B is a pressure versus volume diagram of an Otto cycle of aninternal combustion engine;

FIG. 2A is a pressure versus phase cycle diagram of an Otto cycle of aninternal combustion engine;

FIG. 2B is a pressure versus phase cycle diagram of a cycle of aninternal combustion engine in accordance with the present invention;

FIG. 3 is a pressure versus volume diagram of a cycle of an internalcombustion engine in accordance with the present invention;

FIG. 4A is a side cross-sectional block diagram view of a piston in atop-centered position in a cylinder in accordance with the presentinvention;

FIG. 4B is a side cross-sectional block diagram view of the piston ofFIG. 4A in an expansion stroke moving away from the top of the cylinderand located at compression volume of combustion chamber;

FIG. 4C is a side cross-sectional block diagram view of piston of FIG.4A in a bottommost position in the cylinder;

FIG. 4D is a side cross-sectional block diagram view of the piston inthe cylinder of FIG. 4A compressed and at the point of ignition of fuelin the cylinder;

FIG. 4E is a side cross-sectional block diagram view of the cylinder ofFIG. 4A after a power stroke in response to the ignition depicted inFIG. 4D;

FIG. 4F is a side cross-sectional block diagram view of the cylinder ofFIG. 4A during an exhaust stroke at the point of effective volume;

FIG. 5 is a block diagram view of another embodiment of a cylinder andpiston portion of an engine in accordance with the present invention;

FIG. 6 is a diagram depicting the efficiency of the engine depicted inFIGS. 1A-B and 2A of an Otto cycle engine;

FIG. 7 is a block diagram of another embodiment of portions of an enginein accordance with the present invention;

FIG. 8 is a diagram depicting the relative efficiency of the fueleconomy of the engine depicted in FIGS. 4A-F relative to a conventionalOtto engine; and

FIG. 9 is a block diagram depicting a thermal energy recovery systemused with an engine in accordance with the present invention.

DETAILED DESCRIPTION

In accordance with the principles of the present invention, an internalcombustion engine system and methods for generating power using aninternal combustion engine are provided.

FIG. 3 depicts the conventional Otto cycle (2-4-5-6) in pressure-volumecoordinates as also shown above in FIG. 1B, and further depicts a cycleextension (6-7-8-3-2) for a proposed engine and method for generatingpower in accordance with the present invention. The cycle depicted isdifferent from the conventional Otto cycle and includes the followingsteps:

-   -   1-2: Fuel mixture intake at atmospheric pressure    -   2-3: Adiabatic expansion of mixture to below atmospheric        pressure    -   3-2: Adiabatic compression back to atmospheric pressure    -   2-4: Adiabatic compression to the same pressure as in        conventional Otto cycle    -   4-5: Burning fuel mixture as in conventional Otto cycle    -   5-6-7-8: Adiabatic expansion of gas to the pressure which is        always lower than in Otto cycle (can be even below the        atmospheric pressure)    -   8-7: Adiabatic expansion to the atmospheric pressure    -   7-2-1: Exhaust of gas at atmospheric pressure

In FIG. 3, V_(MIN) indicates a volume of an engine cylinder at maximumcompression for both a conventional prior art engine cycle and theengine cycle of the present invention, V_(C) and V_(MAX) indicatevolumes at maximum expansion, respectively for conventional and proposedengine cycles. All thermo-dynamical parameters (volumes, temperaturesand pressures) on trace 2-4-5-6 are exactly the same for bothconventional and proposed engines. The only difference is an extra areaoutlined by 2-6-7-2, which is responsible for additional useful workperformed by expanding gas in the present invention. Since the workrequired to compress the fuel mixture (area outlined by 1-2-4-1) is thesame for both engines, this will result in increased efficiency of theproposed engine. Thus, an increase of efficiency in the proposed engineis based on the fact that an expansion ratio exceeds a compressionratio, thus using more energy from the power stroke for performinguseful work during an engine cycle.

As depicted in FIG. 3, to achieve the described gain in efficiency, aworking volume of an engine cylinder is increased from V_(C) to V_(MAX).However, it can be easily shown that the same increase in efficiency canbe achieved by a proportional decrease of volume V_(MIN) instead. So itis convenient to consider volumes in FIG. 3 to represent values referredto the volume of maximum compression V_(MIN). Alternatively, the V axisin FIG. 3 could be graduated in values of compression ratio r=V/V_(MIN).In this case, the conventional Otto engine has equal compression andexpansion ratios of V_(C)/V_(MIN), while the proposed engine has acompression ratio of V_(C)/V_(MIN) and an expansion ratio V_(E)/V_(MIN).It is exactly this difference in compression and expansion ratios whichmakes the efficiency of the proposed engine higher.

An example of a cycle of a reciprocating piston 110 in a cylinder 120 ofan engine (not shown) of the present invention as described above isdepicted in FIGS. 4A-F. An intake valve 130 and an exhaust valve 140 maybe controlled by a controller 150 (e.g., a computing unit or otherelectronic or mechanical computing device) driving actuators 131 and 141to control openings and closings of valve 130 and valve 140,respectively. FIG. 4A depicts a beginning of an intake stroke with anair-fuel mixture starting to fill cylinder 120 from volume V_(MIN) upthrough intake valve 130 which is open at a full extension of piston 110toward a top 122 of cylinder 120. While the piston is moving down (i.e.,toward a bottom 160 of cylinder 120) the cylinder is filled to a volumeV_(MAX). No work is done during the process because the mixture is atatmospheric pressure and a constant ambient temperature T₁.

When piston reaches volume V_(C), intake valve 130 closes as depicted inFIG. 4B. The filling process of the air-fuel mixture through intakevalve 130 corresponds to trace 1-2 depicted in FIG. 2A-B as describedabove. After the intake valve closes, a reserve stroke continues tobottom 160 as depicted in FIG. 4C and which corresponds to trace 2-3 inFIG. 3. The mixture expands approximately adiabatically during thereserve stroke to volume V_(MAX) as depicted in FIG. 4C. In a returnstroke, piston 110 returns to compression distance 170 and the mixtureis compressed back to volume V_(C) corresponding to trace 3-2 in FIG. 2.

As indicated, intake valve 130 closes and piston 110 extends through thereserve stroke and the return stroke such that the volume is the samebefore and after the combined reserve stroke and a return stroke. Duringthe reserve stroke, a pressure in a chamber or interior 180 of cylinder120 drops until the piston reaches bottom 160 of the cylinder andincreases during the return stroke as the piston returns to compressionvolume V_(C). Thus, a pressure of the air-fuel mixture in interior 180is lower during the reserve stroke and the return stroke than when a top111 of piston 110 is between the compression distance and top 122 ofcylinder 120.

Also, the reserve stroke and return stroke occur substantiallyadiabatically. Thus, the thermodynamic state of the mixture will besubstantially the same after the piston returns to compression distanceV_(C) after the reserve and return strokes as just before starting theexpansion from V_(C) to V_(MAX) (i.e., the extension of the piston fromthe compression distance to the bottom of the cylinder) in the reservestroke.

Further, a difference in pressure above and below piston 110 during thereserve and return strokes does the work required for moving thecylinder during the reserve and return strokes, which can be eitherpositive or negative, depending on particular pressures. Morespecifically, the work done on trace 3-2 depicted in FIG. 3 will have asame magnitude but the opposite sign compared to the work on trace 2-3,so the total work done on trace 2-3-2 (i.e. the extension of the pistonfrom the compression distance to the bottom of the cylinder and back tothe compression distance) will be equal to zero.

A compression stroke continues from compression volume V_(C) to volumeV_(MIN) as depicted in FIG. 4D. For example, a pressure of the mixturein interior 180 may reach a value equivalent to a conventional engine atV_(MIN). The reserve stroke may be considered to be a portion of anentire intake stroke while the return stroke may be considered to be aportion of an entire compression stroke.

The mixture is then ignited (e.g., using a spark plug) and burned at thetop of the compression stroke, which may cause the pressure andtemperature in interior 180 to rise to values P_(W) and T₃ which may bethe same as in a conventional engine, as depicted at point 4 in FIG. 3.In another example, such a mixture would self-ignite if diesel fuel andno spark plug were used, along with the appropriate pressure forself-ignition of the mixture.

The burned mixture in the interior of the cylinder expands to V_(MAX)during a power stroke depicted in FIG. 4E. For example, interior 180 mayexpand from V_(MIN) to V_(E) shown by trace 5-6-7 in FIG. 3. An interiorpressure in interior 180 at V_(E) may be the ambient pressure of theenvironment around cylinder 120 and any engine of which it is a part.The work contribution to the total cycle work assuming expansion toV_(E) is positive and is determined by formulae described belowinvolving volume change from V_(MIN) to V_(E). Piston 110 may furtherextend such that interior 180 reaches V_(MAX) as depicted in FIG. 4E andFIG. 3 with a bottom 113 of piston 110 extending substantially to bottom160.

V_(MAX) may be a volume of interior 180 at the greatest extent of piston110 relative to top 122. For example, V_(MAX) may be greater than V_(E)as depicted in FIG. 4E. When V_(MAX)≦V_(E), the pressure in interior 180drops to a pressure value above the ambient atmospheric pressure, andthe power stroke concludes at volume V_(MAX)<V_(E).

In an example where V_(MAX) is greater than a volume (i.e, V_(E)) atwhich a pressure of interior 180 would be equal to a pressure of thesurrounding ambient environment (as depicted in FIG. 4E, the pressure ofthe interior drops below V_(E) at an end portion of the power stroke(i.e., from V_(MIN) to V_(MAX)). Specifically, the gas expands to belowthe atmospheric pressure, i.e., to V_(MAX)>V_(E) as shown by trace 7-8in FIG. 3. The work done during the expansion from V_(E) to V_(MAX) isnegative.

After the power stroke depicted in FIG. 4E, piston 110 moves up in abeginning of an exhaust stroke depicted in FIG. 4F with the valves(i.e., valve 130 and valve 140) closed until the volume in interior 180reaches V_(E) again at point 7 in FIG. 3. A substantially adiabaticprocess between V_(E) on the power stroke and V_(E) on the exhauststroke is indicated by trace 7-8-7 in FIG. 3. Due to the reversibilityof this adiabatic process, the gas (i.e., air-fuel mixture) comes to thesame state (e.g., a pressure and temperature state) at point 7 as in thesame point during the power stroke expansion. Similar to trace 2-3-2described above, the total work in trace 7-8-7 is equal to zero,resulting from such substantial adiabatic reversibility.

Upon reaching volume V_(E) at point 7 in FIG. 3, exhaust valve 140depicted in FIG. 4F opens and the burned mixture flows out exhaust valve140 at atmospheric pressure without performing any work. Such exhaustprocess extends until the volume V_(MIN) and is shown by trace 7-2-1 inFIG. 3. The exhaust through value 140 at atmospheric pressurefacilitates the flow of the exhaust from interior 180 of the cylinder tothe surrounding ambient environment.

A total work during the whole cycle depicted in FIGS. 3 and 4A-F iscalculated as a sum of works done on traces 2-3, 3-2, 2-4, 5-7, 7-8,8-7, which is represented by corresponding terms in the followingexpression:

W _(B) =W ₁ −W ₁ −W ₂ +W ₃ +W ₄ −W ₄ =W ₃ −W ₂

where:

W₁ and −W₁ is the work during the reserve (i.e., from V_(C) to V_(MAX))and return strokes (i.e., from V_(MAX) to V_(C))

W₂ is the work done during compression of mixture from V_(C) to V_(MIN)

W₃ is the work done during expansion of combustion gas from V_(MIN) toV_(E)

W₄ and −W₄ is the work during the expansion from V_(E) to V_(MAX) andcontraction from V_(MAX) to V_(E).

This work W_(B) is represented in FIG. 3 by the area outlined by2-4-5-7-2 and is greater than similar work Ws done by a standardconventional engine, which is equal to the area outlined by 2-4-5-6-2.Thus, the effectiveness of the engine of the present invention exceedsthat of a standard engine by a factor:

E _(BS) =W _(B) /W _(S)

In contrast to the conventional Otto cycle, an intake stroke in theengine of the present invention consists of two parts represented bytraces 1-2 and 2-3 in FIG. 3. Trace 1-2 corresponds to filling thecylinder with an air-fuel mixture at atmospheric pressure, which ideallyrequires no work, the same as in a conventional engine. At point 2, theintake valve (e.g., intake valve 130) closes, resulting in an injectionof exactly a substantially same amount of fuel mixture as would beinjected in a conventional engine. Point 2 depicted in FIG. 3 is a point(e.g., compression volume V_(C)) at which a compression of the fuel airmixture (i.e, from V_(C) to V_(MIN)) would yield a same conventionalcompression ratio maximized for a given fuel type (i.e., while avoidingdetonation). However, in the engine depicted in FIGS. 4A-F, theexpansion continues substantially adiabatically to working volumeV_(MAX) (via the reserve stroke described above) in accordance with theadiabatic equation:

PV^(γ) = Const = K${{where}\mspace{14mu} \gamma} = \frac{C_{P}}{C_{V}}$

where C_(P) and C_(V) are constant pressure and constant volume heatcapacities of the gas mixture (mainly air).

The work in the adiabatic process can be calculated as

W = ∫_(V_(i))^(V_(f))P V so  that$W = {K{\int_{V_{i}}^{V_{f}}\ {\frac{V}{V^{\gamma}}\mspace{14mu} {or}}}}$$W = \frac{K\left( {V_{f}^{1 - \gamma} - V_{i}^{1 - y}} \right)}{1 - \gamma}$

As described above, the compression stroke includes two partsrepresented by trace 3-2 (also described as a return stroke) and trace2-4, which are also substantially adiabatic. Taking into considerationthe (approximate) reversibility of adiabatic processes as implemented inICE, the work done depends only on initial and final pressures andvolumes. Thus, the total work done in trace 2-3-2 is substantially equalto zero. As a result at the end of trace 3-2, the fuel mixture comes tosubstantially the same state as in the beginning of the compressionstroke of the conventional engine.

Thus, it is always possible at any given constructional compressionratio (i.e, V_(MAX)/V_(MIN)) to fill a cylinder (e.g., cylinder 120)with an amount of a fuel-air mixture necessary for optimal operation(i.e., the maximum amount which still avoids detonation) for aconventional engine with compression ratio V_(C)/V_(MIN).

For example, for a given amount of fuel mixture, if it is considered tobe an ideal gas, all equilibrium processes, including the adiabaticprocess obeys the ideal gas equation:

$\frac{P_{atm}V_{c}}{T_{1}} = \frac{P_{C}V_{MIN}}{T_{2}}$

where P_(atm), T₁ are pressure and temperature at the beginning of thecompression (point 2 in FIG. 3) and P_(C), T₂ are pressure andtemperature at the end of the compression (point 4 in FIG. 3).

After compression to volume V_(MIN), the state (e.g., temperature andpressure in an interior of the cylinder) of the mixture will be the sameregardless of when or how during the intake stroke the mixture hasfilled the cylinder. Consequently, it is not necessary to fill thecylinder along the trace 1-2-3-2 in FIG. 2. It is possible to fill thecylinder slower, or intermittently, or even during the process 3-2. Itis only important to achieve the same state (e.g., temperature andpressure) of the mixture at the end of the process 3-2. Therefore, thework required for compression of the fuel mixture depends only oninitial and final thermo-dynamical states of the mixture and will notdepend on a particular time of filling of the cylinder.

The power stroke is a standard adiabatic expansion shown by the trace5-6-7-8 in FIG. 3. The work released depends only on the final andinitial volumes and is given by:

$W = \frac{K\left( {V_{f}^{1 - \gamma} - V_{\iota}^{1 - y}} \right)}{1 - \gamma}$

In contrast to the work required for compression of a mixture, which isthe same for the cylinder described above relative to FIGS. 3 and 4A-Fas in a conventional engine, the work during the power stroke (i.e.,from V_(MIN) to V_(MAX) in FIGS. 4D-E and trace 5-6-7-8) exceeds thework done for a conventional engine. As depicted in FIG. 3, the areaoutlined by the trace 2-4-5-6-2 represents the useful work done in theconventional Otto cycle. The area outlined by 2-6-7-2 is an additionalwork done by the proposed engine described and depicted herein, with theadditional work resulting in increased efficiency relative to aconventional Otto engine.

As is evident from FIG. 3, the extra area (i.e., the area outlined by2-6-7-2) results from V_(MAX) being always greater than V_(C) in theengine described and depicted in FIGS. 3 and 4A-F. Thus, a physicalcompression ratio V_(C)/V_(MIN) is always less than the constructionalcompression ratio V_(MAX)/V_(MIN). An ideal value of V_(MAX) would beV_(MAX)=V_(E), where V_(E) is the volume at point 7 where pressure dropsdown to the atmospheric value P_(atm) as described above. V_(E) providesan advantageous condition for exhaust of the results of the combustionsince such exhaust occurs at atmospheric pressure.

However, when V_(MAX)>V_(E) the expansion during the expansion strokedescribed above continues to point 8 in FIG. 3 where the pressureP<P_(atm). In another example, the volume of interior 180 may beadjusted to maintain optimal combustion conditions for fuels withdifferent octane ratings. In particular, engine geometry (e.g., a sizeof the cylinders including maximum and minimum volumes thereof) may beoptimized for a fuel with a highest octane rating that is anticipated tobe used in the engine such that V_(MAX)=V_(E) for this given fuel. Toavoid detonation in the same engine when using lower octane fuels, V_(C)is reduced to capture less air-fuel mixture and to decrease a workingpressure P_(W). This causes the pressure at point 8 (FIG. 3) to dropbelow the atmospheric level at volumes greater than V_(E). However,since both the expansion 7-8 and the compression 8-7 are approximatelyreversible adiabatic processes as described above, the total work on thetrace 7-8-7 is about zero, and thus does not affect the efficiency ofthe engine. Instead, the efficiency is automatically maintained at themaximum possible level for any octane rating of the fuel and can becalculated as if the gas always expands to an effective volume V_(E).Thus, a volume (V_(C)) of the cylinder and therefore a compressiondistance for a compression stroke may be adjusted based on the octanelevel of a fuel to be used. For example, a user may indicate to acontroller the fuel which is being used, or a sensor may determine suchfuel, such that an intake valve may be opened and closed by an actuatorcontrolled by the controller to allow a fuel—air mixture along with acompression distance to be adjusted based on the type of fuel used.

Turning to FIGS. 2A-2B, in conventional engine design a maximumcompression of intake gas and expansion ratios is the same, being equalto the ratio of maximum to minimum cylinder volumes, r. A consequence ofthis ratio equality is that the temperature changes by the same factorduring adiabatic compression and expansion processes. The adiabaticequation relating temperature and volume TV^(γ-1)=Constant implies thisif the volume is changing by the same factor r in both processes, so isthe temperature:

T ₃ /T ₄ =T ₂ /T ₁

where T₁, T₂, T₃, T₄ are gas temperatures in points 1, 2, 3, 4 of theOtto cycle shown in FIG. 2A.

Relative to the engine described above relative to FIGS. 3 and 4A-F,FIG. 2B depicts a cycle diagram different than that for a prior artengine depicted in FIG. 2A. Both the intake and exhaust strokes areimportant for the engine thermodynamics in FIG. 2B. Both theconventional Otto cycle (FIG. 2A) and the proposed cycle (FIG. 2B) ofthe present invention are shown for a maximum filling of cylinders withthe air-fuel mixture. The cycle in FIG. 2B always realizes the sametemperatures and pressures at points 2 and 3 as in conventional Ottocycle depicted in FIG. 2A. However, the temperatures and pressures atpoints 1 and 4 are always lower in the engine described herein relativeto the conventional Otto cycle of FIG. 2A. Moreover, the temperature andpressure at point 4 may also go even lower than atmospheric values. FIG.3 also depicts same conventional Otto cycle (i.e., trace 1-2-4-5-6-2) ofFIG. 2A along with a cycle extension (i.e., trace 6-7-8-3-2) for theengine described herein as described above.

Using the notations of FIGS. 2A-2B for a conventional Otto cycle, theefficiency q of an engine can be expressed as follows:

$\begin{matrix}{\eta = {\frac{W_{net}}{g_{in}} = {{1 - \frac{g_{out}}{g_{in}}} = {1 - \frac{T_{4} - T_{1}}{T_{3} - T_{2}}}}}} & (1)\end{matrix}$

The fact that compression and expansion ratios are equal in aconventional engine leads to the following relation:

$\begin{matrix}{\frac{T_{3}}{T_{4}} = {\frac{T_{2}}{T_{1}} = {\left( \frac{V_{C}}{V_{MIN}} \right)^{k - 1} = r^{k - 1}}}} & (2)\end{matrix}$

where r is a compression ratio for the conventional Otto cycle, and k isan effective value of γ from formula (3) which is equal to 1.35 fortypical air-fuel mixture. From (2) it follows that:

T₃=T₄r^(k-1)  (3)

T₂=T₁r^(k-1)  (4)

T ₃ −T ₂=(T ₄ −T ₁)r ^(k-1)  (5)

From (1) and (5) we have:

$\begin{matrix}{\eta = {1 - \frac{1}{r^{k - 1}}}} & (6)\end{matrix}$

FIG. 6 depicts a classical derivation based on the proportional relationbetween a change in internal energy and a change in temperature and canbe found in numerous classical textbooks as well as in recentpublications, e.g., the derivation depicted was taken from the followingreference: http://www.engr.colostate.edu/˜allan/thermo/page5/page5.html

For the purpose of efficiency calculations in the engine describedherein relative to FIGS. 3 and 4A-F, it must be taken into considerationthat the compression and expansion ratios are different. According toFIG. 3, the gas is compressed above atmospheric pressure from volumeV_(C) to V_(MIN) (i.e., trace 2-4), but expands from volume V_(MIN) toeffective volume V_(E) (i.e., trace 5-6-7). To account for this change,the formula (I) must be modified as follows:

$\begin{matrix}{\eta = {1 - \frac{T_{E} - T_{1}}{T_{3} - T_{2}}}} & (7)\end{matrix}$

The relation between the temperatures and volumes during compressiongives:

$\begin{matrix}{\frac{T_{2}}{T_{1}} = {\left( \frac{V_{C}}{V_{MIN}} \right)^{k - 1} = r_{C}^{k - 1}}} & (8)\end{matrix}$

where r_(C) is the compression ratio for the proposed engine.

The similar relation for expansion gives:

$\begin{matrix}{\frac{T_{3}}{T_{E}} = {\left( \frac{V_{E}}{V_{MIN}} \right)^{k - 1} = r_{E}^{k - 1}}} & (9)\end{matrix}$

where r_(E) is the expansion ratio for the proposed engine. ExpressingT₂ and T_(E) from (8) and (9) and inserting them into (7) gives thefollowing expression for η:

$\begin{matrix}{\eta = {{1 - \frac{\frac{T_{3}}{r_{E}^{k - 1}} - T_{1}}{T_{3} - {T_{1} \cdot r_{C}^{k - 1}}}} = {1 - \frac{{\left( {T_{3}/T_{1}} \right)/r_{E}^{k - 1}} - 1}{{T_{3}/T_{1}} - r_{C}^{k - 1}}}}} & (10)\end{matrix}$

Note that η now depends on the ratio T₃/T₁, which has an approximatevalue of 8 for most typical gasoline engines. Assuming the compressionratio of the best gasoline engines r_(C)≈10 the doubled expansion ratioof the proposed engine, r_(E)≈20, the value of η can be calculated as:

$\begin{matrix}\begin{matrix}{\eta = {1 - \frac{{8/r_{E}^{k - 1}} - 1}{8 - r_{C}^{k - 1}}}} \\{= {1 - \frac{{8/2.85} - 1}{8 - 2.24}}} \\{= {1 - \frac{1.8}{5.76}}} \\{= {1 - 0.31}} \\{= 0.69}\end{matrix} & (11)\end{matrix}$

The same formula (10) can be used to calculate η for a conventionalengine, which gives:

$\begin{matrix}{\eta = {{1 - \frac{{8/2.24} - 1}{8 - 2.24}} = 0.55}} & (12)\end{matrix}$

Note that this value is in a perfect agreement with a classical formula(14):

$\begin{matrix}{\eta = {{1 - \frac{1}{r^{k - 1}}} = {{1 - \frac{1}{10^{0.35}}} = {{1 - \frac{1}{2.24}} = {{1 - 0.45} = 0.55}}}}} & (13)\end{matrix}$

As will be understood by one skilled in the art, the listed algorithmsmay be performed on a computing unit or controller and may be stored oncomputer readable storage mediums which may be read and executed by sucha computing unit or controller (e.g., controller 50, controller 235,controller 335). FIG. 8 is a diagram depicting the relative efficiencyof the fuel economy of the engine depicted in FIGS. 4A-F relative to aconventional Otto engine with a compression ratio of 8.

In another example depicted in FIG. 5, the cylinder and piston describedabove relative to FIGS. 3 and 4A-F may be modified to include anadditional intake mechanism or valve upstream of a standard intakevalve. An engine (not shown) including a cylinder 220 and piston 210 isconfigured with a constructional compression ratio (i.e.,V_(MAX)/V_(MIN)) exceeding a compression ratio (e.g., up to about 11.5)of a standard gasoline engine, and even exceeding this ratio in a dieselengine (e.g., up to about 25). A main intake valve 230 is a standardmechanical valve as in a conventional engine made of high strength steelto allow the valve to hold high pressure during combustion. As in aconventional engine, intake valve 230 opens in a beginning of an intakestroke of a piston 210 in a cylinder 220 and closes at the end of such astroke. A second valve 231 is installed just before (i.e., upstream) ofvalve 230 and is used for adjusting an amount of an incoming gasmixture. Valve 231 opens and closes by means of an actuator 232controlled by a controller 235 (e.g., a computing unit).

Controller 235 receives data about an angular speed w of a crankshaft250, a temperature of the air-fuel mixture T₁ and/or a crankshaftposition which determines a current cylinder volume V_(C) via one ormore sensors 237 configured to determine such parameters. Also, pressuresensors may determine an interior pressure of interior 210 which may beprovided to controller 235. Controller 235 also receives or obtains dataabout parameters of the gas-fuel mixture, including the effective valueof k.

Controller 235 determines a time for opening and closing of valve 231,which may be performed according to the following algorithm:

-   -   a. First, a maximum possible V_(C), at which the fuel mixture        will burn without detonation, is computed by the formula:

$\begin{matrix}{V_{C\_ max} = \frac{C_{F}V_{MIN}T_{1}}{P_{1}}} & (14)\end{matrix}$

-   -   -   where C_(F) is the factor determined by critical pressure            P_(C) and temperature T_(C), at which the mixture is close            to detonation, as follows:

$\begin{matrix}{C_{F} = \frac{P_{C}}{T_{C}}} & (15)\end{matrix}$

-   -   -   where T_(C)=T₂. The factor C_(F) is measured experimentally            for each particular engine design and fuel type. The            measurements should be carried out at low angular velocity w            of the engine, at which the resistance of intake manifold to            the mixture flow doesn't depend on w.

    -   b. At higher values of w, V_(C) _(—) _(max) has to be adjusted        upward due to higher resistance to the mixture flow. The        behavior of V_(C) _(—) _(max) as a function of w is measured        experimentally for every particular engine design.

    -   c. The actual value of V_(C) is then computed by the formula:

V _(C) =F·V _(C) _(—) _(max)  (16)

-   -    where F is the position factor of the accelerator pedal ranging        from 0 to 1.

Controller 235 causes actuator 232 to open valve 231 in the beginning ofintake stroke and closes it when the volume reaches the computed valueV_(C), Controller 235 may also open and close valve 231 at other momentsof time during the intake stroke, based on required engine power. Suchopening and closing may be done to avoid a maximum amount of mixturedrawn into an interior 280 of cylinder 220 exceeding an amount causingdetonation.

Resulting from the use of controller 235 to control second intake valve231 an optimal value of C_(F) for the engine operation at maximum powerper cycle without detonation may be maintained automatically. The valueof C_(F) may be maintained at a constant level independent of:

-   -   a. Temperature variations of the intake fuel mixture    -   b. Variations in pressure at the intake manifold, and        consequently, independent of a change in flow resistance of air        filters.    -   c. Variations in design of the intake manifold, which results in        different air flow resistance. The requirement of minimization        of the intake manifold resistance, which is vitally important in        conventional engines, is also unnecessary.

Further, the engine cycle power may be maintained at a constant level,set by the accelerator pedal, independently of the parameters listedabove in paragraphs a-c. The engine may use the fuel (e.g., gasoline,ethanol, organic gases, biodiesel, etc.) of any octane rating includingdiesel fuel, with almost the same efficiency. A change in octane ratingwill only cause a change in the cycle power, because any required fuelmixture can be accommodated by small adjustment of V_(C). Morespecifically, such adjustment of V_(C) results in only minor variationsin compression ratio r_(C) and consequently, the efficiency calculatedby the formula (10) described above. For diesel fuel engines, therequired V_(C) is almost twice as high as that for gasoline engines and,to avoid a noticeable drop in power, the constructional compressionratio (i.e, V_(MAX)/V_(MIN)) must be correspondingly increased.

In another example, a controlled damper may be utilized instead of anadditional valve as shown in FIG. 7. A regulated damper 332 is installedat the input of an intake collector or receiver 334, which is common forall cylinders of an engine (not shown) of which a cylinder 320 is apart. A position of the damper position determines an amount of mixtureentering the cylinders by creating a resistance to a mixture flow andthus controlling the average pressure P_(x) and temperature T_(x) insideof the receiver. An intake valve 330 may be kept open for an entireintake stroke (e.g., including a reserve stroke as described above) asin a conventional engine and thus may be controlled by the same simplemechanical methods with fixed timing as may be done in a conventionalengine. For example, the receiver may be in fluid communication with theinteriors of multiple cylinders via multiple such intake valves (e.g.,intake valve 330). An interior of the receiver may include thermalinsulation to regulate a temperature and flow rate of the fuel-airmixture from the damper toward the intake valve(s).

Due to reversibility of the substantially adiabatic process describedabove, the engine operation is independent of a specific timing at whichthe mixture fills the cylinder, as long as a molar amount (e.g., a mass)of a taken mixture received in an interior 380 of cylinder 320 is thesame. As described above relative to FIG. 5, the maintenance of such amolar amount is satisfied by setting V_(C) according to the expression(14) for a given fuel coefficient C_(F) determined by (15). In case ofregulated damper 332, a same required amount of the fuel mixture may beprovided by adjusting the damper position so that P_(x) and T_(x)satisfy the following relation:

$\begin{matrix}{\frac{P_{x}V_{MAX}}{T_{x}} = {C_{F}V_{MIN}}} & (17)\end{matrix}$

This equation (17) above assumes that the air resistance of the intakevalve is negligible, so that the pressure and temperature at point 3 inFIG. 3 are approximately equal to P_(x) and T_(x) measured by a sensor(e.g., one of sensors 337) inside the receiver. In a real situation, theamount of mixture will depend on the engine RPM (i.e., speed of thecrankshaft, pistons, etc. in revolutions per minute) because of theincrease of the intake valve resistance with increased RPM's. To accountfor this effect the equation (17) must be rewritten as follows:

$\begin{matrix}{\frac{P_{x}V_{MAX}}{T_{x}} = {K_{rpm}C_{F}V_{MIN}}} & (18)\end{matrix}$

where K_(rpm) is a correction coefficient which is experimentallymeasured as a function of RPM for each particular engine design.

A controller 335 (e.g, a computing unit) may maintain an optimal fillingof cylinders by controlling a position of damper 332 so that themeasured values of P_(x) and T_(x) satisfy the relation (18). Controller335 may also receive data about an angular speed w of a crankshaft 350,a temperature of the air-fuel mixture T₁, and/or a crankshaft positionwhich determines a current cylinder volume V_(C) via one or more sensors337 configured to determine such parameters. Also, a pressure sensor maydetermine an interior pressure of interior 280 which may be provided tocontroller 335.

The damper may utilize a more sophisticated computing algorithm relativeto the additional valve (i.e., valve 231) described above, and the useof the damper significantly simplifies the mechanical design relative tothe additional valve. A damper is particularly advantageous for amulti-cylinder engine, because only one damper may be used for the wholeengine. This possibility results from the fact that the sequentialoperation of cylinders in a multi-cylinder engine causes a steadynon-pulsating reduction of the pressure in a receiver (e.g., receiver334) below the atmospheric level. The pressure and temperaturemeasurements are then averaged by the receiver and may be used tocontrol a position of the damper, which is common for all cylinders.

Returning to FIGS. 2-3 and the cycle depicted in FIGS. 4A-F, a pressurein an interior of a cylinder (e.g., cylinder 120) during a power strokemay drop below an atmospheric level, crossing it at the point 7 of anexpansion trace 5-6-7-8, where V=V_(E). This point corresponds to themaximum expansion ratio r_(E), at which the burned mixture stillperforms positive work, so that r_(E) is used in the formula (8) tocalculate efficiency, even in the case of V_(MAX)>V_(E). As previouslydiscussed, it may be advantageous to utilize an engine having cylindersand pistons configured with V_(MAX)>V_(E) to enable using fuels with awide range of octane ratings. For this purpose, the engine must bedesigned so that V_(E)=V_(MAX) for the highest octane rating to beutilized with the engine. For lower octane ratings, V_(E) is thusreduced to V_(E)<V_(MAX) and the pressure drops below atmospheric whenV>V_(E).

To prevent atmospheric air from entering the cylinder in this case, theexhaust valve (e.g., valve 140 in FIGS. 4A-F) must open when volumereaches V_(E) on the trace 8-7 (FIG. 3). Alternatively, such opening maybe carried out by an additional low-pressure exhaust valve (not shown),which would open when the pressure reaches the atmospheric level atpoint 7 of trace 8-7.

In many prior art designs it would be reasonable to maintainV_(MAX)≦V_(E) to avoid a decrease of a cycle power due to increases inRPM of a corresponding engine cycle. However, in the engine disclosedherein, it is possible to maintain the cycle power at constant level fordifferent engine RPMs, because such an increase of an intake manifold'sresistance due to increased RPMs may be compensated for by adjusting anopening time of valve 130 depicted in FIG. 4. More particularly, theadjusting of such an opening and closing of the intake valve maintains asame amount of fuel mixture at different RPMs, consequently resulting ina constant cycle power. Preferably, a required volume V_(C) of a fuelmixture may be provided by controlling a closing time of the intakevalve (e.g., valve 130 depicted in FIG. 4). However, the same result maybe achieved by adjusting a resistance of an intake manifold of theengine, for example with a damper (e.g., damper 332 depicted in FIG. 7),so that an amount of mixture provided to an interior (e.g., interior180) of a cylinder of an engine is substantially an amount desired tomaintain a desired cycle power at any RPM. For example, damper 332 maybe controlled by controller 335 to adjust an amount of a fuel-airmixture flowing into cylinder 320 based on an octane rating of fuelwhich is utilized.

Further the requirements for the design of intake manifold and airfilters may be significantly simplified for the described engine. Airfilters may be made from denser materials and their lifetime can besignificantly extended because the amount of the mixture may beautomatically maintained (e.g., by a controller) by adjusting a damperposition (e.g., of damper 332) or a timing of an additional intake valve(e.g., intake valve 231 depicted in FIG. 5) in accordance with pressurevariations in the intake manifold. Further, the requirements for exhaustvalves may be also simplified because the exhaust occurs at a much lowerpressure which is close to, or at about, an atmospheric level.

Although, the main anticipated application of the engine describedherein is the automotive industry, it is not intended to limit use insuch a manner described. All gasoline engines may utilize the advantagesof the proposed principle, which will, on average, save about 25% offuel for a same cycle power. A resulting drop in the cycle power can becompensated by increased engine RPM, taking advantage of the describedengine's ability to maintain constant amount of fuel mixture in a widerange of engine RPMs (i.e., via opening and closing valve 130, valve231, and damper 332).

In one example, the engine described above may be used in electric powergenerators including engines in hybrid cars. In such a case, thedescribed engine may be used for supplying electric power for the cardrive train or charging batteries of the car when the car isn't moving.The described engine in this application may be optimized for a fixedRPM yielding the maximum possible efficiency, which will additionallysimplify the overall design. Another example is in stationary powergenerators, where the consideration of engine size is secondary to therequirement of having the best possible efficiency.

As described above, operation of the engine described herein wouldresult in exhaust gas at about atmospheric pressure and having a hightemperature of about 1300 K. In conventional engines, the exhaust gashas a much higher temperature of about 1500-1800 K and pressure of about4-6 atm, which requires use of comprehensive exhaust manifolds anddampers of acoustic vibrations. The relatively low exhaust pressure(e.g., close to atmospheric level) makes it possible to significantlysimplify the exhaust manifold construction and avoid or minimize use ofacoustic dampers. Further the exhaust heat may be recaptured via use ofa heat exchanger, which may be utilized, for example, for rapid heatingof a car interior during the winter. In another example, the exhaustheat may be converted into electricity, for example by a steam powergenerator which can provide additional power for charging a car battery.

Further, in another example, a system for recapturing heat from theexhaust and from the engine coolant utilized in the example describedabove is depicted in FIG. 9. Internal combustion engine coolant flowsthrough a steam condenser 400. A fluid pressure pump 410 pumps fluidfrom the steam condenser to a heat exchanger or fluid steamer whichreceives heat from the exhaust gases of an engine in accordance with oneof the examples described above. Steam flows from the heat exchanger orfluid steamer to a steam powered generator which generates mechanical orelectrical power. Low pressure steam then flows to the steam condenserwhich also acts as a heat exchanger with the internal combustion enginecoolant flow.

Further, a proper utilization of the thermal energy of the steam cooleddown from 1300 K to 400 K may save almost half of energy normally lostby an engine via the exhaust gas. This would result in the increase ofthe overall efficiency of the described engine to about 80-85%, which isnoticeably higher than that of a conventional engine, which typicallydoesn't exceed 50%. Such a steam power generator may be of aclosed-cycle type and may use liquids with smaller heat of evaporation,similar to liquids used in air conditioning systems.

Further, it is important to note that the described engine consumesalmost half as much fuel per one cycle relative to a standard enginewhile the described engine causes a decrease of the cycle power by only15-20%. This fact, together with a direct proportionality between thetotal engine power and its RPM, makes it possible to create engines,which will consume half as much fuel at a same cylinder volume. Thepower output per unit volume of the cylinder, and also the power perunit mass of the engine described herein, will outperform prior artengines, even without utilizing the steam generator described above.

Further, as will be understood from the above the described engineallows a more efficient conversion of the internal thermal energy ofexpanding burned fuel mixture into mechanical energy. As described, gascompression and expansion in the engine occurs according to anapproximately reversible adiabatic processes. Moreover, this approximatereversibility is used to allow the manipulating the rate of intake ofthe fuel mixture, thereby making it possible to significantly increasethe expansion ratio (e.g., V_(MAX)/V_(MIN)) without a correspondingincrease of the final pre-ignition compression pressure in the cylinder(e.g., as compared to a conventional engine), thus avoiding undesirabledetonation.

Although preferred embodiments have been depicted and described indetail herein, it will be apparent to those skilled in the relevant artthat various modifications, additions, substitutions and the like can bemade without departing from the spirit of the invention and these aretherefore considered to be within the scope of the invention as definedin the following claims.

1. A method for use in generating power in an internal combustion engine comprising: controlling a flow of a gas having an initial temperature and an initial pressure to a combustion chamber of the engine by a controller through an intake mechanism to provide a first mass of the gas to the chamber; expanding the combustion chamber from a minimum volume to a maximum volume in an intake stroke, the maximum volume of the chamber exceeding a maximum compression volume of the chamber, the first mass of the gas in the chamber having a first pressure and a first temperature at the maximum compression volume; reducing the chamber to the compression volume from the maximum volume in a compression stroke, the compression volume being about a volume of the chamber such that the first mass of the gas at the first pressure and the first temperature in the chamber is less than a maximum mass detonating with a fuel by the end of the compression stroke; wherein the expanding the chamber from the compression volume to the maximum volume during the intake stroke and the compressing the chamber to the compression volume during the compression stroke comprises a process which is about adiabatic and about reversible and wherein the gas has a second pressure and a second temperature during the process lower than the initial temperature and the initial pressure; igniting a mixture of the gas and the fuel in the chamber to expand the chamber from the minimum volume to the maximum volume in a power stroke; reducing the chamber to the minimum volume from the maximum volume in an exhaust stroke and exhausting exhaust gas and burned fuel resulting from the igniting into a surrounding ambient environment; and expanding the chamber from the minimum volume to the maximum volume in a next intake stroke, and controlling the flow of the gas to provide the first mass of the gas to the chamber during the next intake stroke.
 2. The method of claim 1 wherein the exhaust gas and the burned fuel are exhausted during the compression stroke when a pressure of the exhaust gas and the burned fuel in the chamber is about equal to a pressure of the surrounding ambient environment.
 3. The method of claim 1 wherein the maximum volume is greater than or equal to an effective expansion volume, the effective expansion volume comprising a volume when a pressure of the exhaust gas and the burned fuel in the chamber is equal to a pressure of the surrounding ambient environment.
 4. The method of claim 3 further comprises controlling the exhausting of the exhaust gas out of the combustion chamber by a controller via exhaust mechanisms at the effective expansion volume.
 5. The method of claim 1 wherein the controlling the flow of the gas comprises the controller controlling the flow of the gas based on the initial temperature and the initial pressure.
 6. The method of claim 1 wherein the controlling the flow comprises the controller controlling the flow of the gas based on an angular speed of a crankshaft of the engine.
 7. The method of claim 1 wherein the controlling the flow comprises the controller controlling the flow of the gas based on a crankshaft position determining an instantaneous volume of the combustion chamber during the intake stroke.
 8. The method of claim 1 wherein the controlling the flow comprises the controller controlling the flow of the gas based on a type of the fuel.
 9. The method of claim 1 wherein the controlling the flow comprises the controller controlling the flow of the gas based on a speed of a change in a volume of the combustion chamber.
 10. The method of claim 1 wherein the controlling the flow comprises the controller controlling the flow of the gas based on a resistance to a flow of the gas through an intake manifold toward the chamber.
 11. The method of claim 1 further comprising coupling a heat exchanger to an exhaust manifold receiving the exhaust, the heat exchanger coupled to a steam generator configured to generate additional power.
 12. The method of claim 1 wherein the intake mechanism comprises an intake valve controlled by the controller.
 13. The method of claim 1 wherein the combustion chamber is bounded by a cylinder of the engine and the intake mechanism comprises a low pressure intake valve controlled by the controller and upstream of a high pressure intake valve of the cylinder, the high pressure intake valve allowing fluid communication between the chamber and an exterior of the chamber, the high pressure intake valve configured to remain closed during the power stroke and configured to remain closed at a higher pressure than the low pressure intake valve is configured to remain closed.
 14. The method of claim 1 wherein the intake mechanism comprises a damper controlled by the controller.
 15. The method of claim 1 wherein the fuel comprises a fuel ignitable by a spark plug.
 16. A method for use in generating power in an internal combustion engine comprising: controlling a flow of a gas having an initial temperature and an initial pressure to a combustion chamber of the engine by a controller through an intake mechanism to provide a first mass of the gas to the chamber; expanding the combustion chamber from a minimum volume to a maximum volume in an intake stroke, the maximum volume of the chamber exceeding a maximum compression volume of the chamber, the first mass of the gas in the chamber having a first pressure and a first temperature at the maximum compression volume; reducing the chamber to the compression volume from the maximum volume in a compression stroke, the compression volume being about a volume of the chamber such that the first mass of the gas at the first pressure and the first temperature in the chamber is more than a minimum mass self-igniting with a fuel by the end of the compression stroke; wherein the expanding the chamber from the compression volume to the maximum volume during the intake stroke and the compressing the chamber to the compression volume during the compression stroke comprises a process which is about adiabatic and about reversible and wherein the gas has a second pressure and a second temperature during the process lower than the initial temperature and the initial pressure; igniting a mixture of the gas and the fuel in the chamber to expand the chamber from the minimum volume to the maximum volume in a power stroke; reducing the chamber to the minimum volume from the maximum volume in an exhaust stroke and exhausting exhaust gas and burned fuel resulting from the igniting into a surrounding ambient environment; and expanding the chamber from the minimum volume to the maximum volume in a next intake stroke, and controlling the flow of the gas to provide the first mass of the gas to the chamber during the next intake stroke.
 17. The method of claim 16 wherein the exhaust gas and the burned fuel are exhausted during the compression stroke when a pressure of the exhaust gas and the burned fuel in the chamber is about equal to a pressure of the surrounding ambient environment.
 18. The method of claim 16 wherein the maximum volume is greater than or equal to an effective expansion volume, the effective expansion volume comprising a volume when a pressure of the exhaust gas and the burned fuel in the chamber is equal to a pressure of the surrounding ambient environment.
 19. The method of claim 18 further comprises controlling the exhausting of the exhaust gas out of the combustion chamber by a controller via exhaust mechanisms at the effective expansion volume.
 20. The method of claim 16 wherein the controlling the flow of the gas comprises the controller controlling the flow of the gas based on the initial temperature and the initial pressure.
 21. The method of claim 16 wherein the controlling the flow comprises the controller controlling the flow of the gas based on an angular speed of a crankshaft of the engine.
 22. The method of claim 16 wherein the controlling the flow comprises the controller controlling the flow of the gas based on a crankshaft position determining an instantaneous volume of the combustion chamber during the intake stroke.
 23. The method of claim 16 wherein the controlling the flow comprises the controller controlling the flow of the gas based on a type of the fuel.
 24. The method of claim 16 wherein the controlling the flow comprises the controller controlling the flow of the gas based on a speed of a change in a volume of the combustion chamber.
 25. The method of claim 16 wherein the controlling the flow comprises the controller controlling the flow of the gas based on a resistance to a flow of the gas through an intake manifold toward the chamber.
 26. The method of claim 16 further comprising coupling a heat exchanger to an exhaust manifold receiving the exhaust, the heat exchanger coupled to a steam generator configured to generate additional power.
 27. The method of claim 16 wherein the intake mechanism comprises an intake valve controlled by the controller.
 28. The method of claim 16 wherein the combustion chamber is bounded by a cylinder of the engine and the intake mechanism comprises a low pressure intake valve controlled by the controller and upstream of a high pressure intake valve of the cylinder, the high pressure intake valve allowing fluid communication between the chamber and an exterior of the chamber, the high pressure intake valve configured to remain closed during the power stroke and configured to remain closed at a higher pressure than the low pressure intake valve is configured to remain closed.
 29. The method of claim 16 wherein the intake mechanism comprises a damper controlled by the controller.
 30. The method of claim 16 wherein the fuel comprises a self-ignitable fuel. 